Hot water supply system

ABSTRACT

A hot water supply system wherein a flow of a refrigerant on a high-pressure side of a supercritical heat pump cycle and a flow of hot water are oppositely directed, and wherein water heated by the refrigerant on the high-pressure side is stored in heat insulating tanks of a vacuous and double structure. Water is heated with heat absorbed from the atmosphere and by the supercritical heat pump cycle having a high heat exchange efficiency η, to reduce the power required to heat the water. Therefore, hot water can be generated even in the daytime during peak power rates. Accordingly, because it is unnecessary to store hot water for daytime use in insulated tanks, system space requirements are reduced.

CROSS-REFERENCE TO RELATED APPLICATIONS

The present application is related to, and claims priority from,Japanese Patent Application Nos. Hei. 10-328538 and 11-264336, thecontents of which are incorporated herein by reference.

BACKGROUND OF THE INVENTION

1. Field of the Invention

This invention relates generally to a hot water supply system, andparticularly to a system that heat exchanges water with a heatedrefrigerant in a heat exchanger while maintaining a difference intemperature between refrigerant exhausted from the heat exchanger andwater input into the heat exchanger within a predetermined temperaturerange.

2. Related Art

A typical hot water supply system is described in, for example, JapaneseUtility Model Laid-Open No. Hei. 6-73652. In the described system, wateris heated by an electric heater and is then stored in a heat insulatingtank. As the water must be heated by the electric heater, overall systempower requirements are increased. Therefore, such a hot water supplysystem generates hot water at night to minimize electricity costs.

SUMMARY OF THE INVENTION

Therefore, it is an object of the present invention to provide a hotwater supply system in which the amount of power required to generatehot water is reduced.

The present invention provides a heating system in which a refrigerantflow on a high-pressure side of a heat pump cycle and a hot water floware opposite one another. Water heated by the high-pressure siderefrigerant is stored in heat insulating tanks.

In the above-described arrangement, water is heated by heat absorbedfrom the atmosphere and by a supercritical heat pump cycle of high heatexchange efficiency, so that power required to heat the water can beminimized. Therefore, it is possible to generate hot water by using aminimum amount of power, as hot water may be generated when demand sorequires, without the need for storing hot water for daytime use.Moreover, as heat insulating tanks are not required for storing waterheated during nighttime hours for daytime use, a much smaller-scale hotwater supply system can be implemented.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a perspective view of a first embodiment of the hot watersupply system according to the present invention;

FIG. 2 is a schematic diagram of the system shown in FIG. 1;

FIG. 3 is a sectional view taken along line III—III in FIG. 2;

FIGS. 4A-4C are schematic diagrams showing operation of the firstembodiment;

FIG. 5A is top plan cross-sectional view of a counterflow-type heatexchanger used in the first embodiment;

FIG. 5B is a graph showing temperature variation versus flow distance inthe counterflow type heat exchanger;

FIG. 6 is a graph showing temperature variation in a water heatexchanger using freon as a refrigerant;

FIG. 7 is a graph of T(temperature) vs. S(specific entropy) forrefrigerant used in the system of the present invention;

FIG. 8 is another graph of T vs. S;

FIG. 9 is another graph of T vs. S;

FIG. 10 is another graph of T vs. S;

FIG. 11 is another graph of T vs. S;

FIG. 12 is another graph of T vs. S;

FIG. 13 is another graph of T vs. S;

FIG. 14 is a schematic diagram showing a heat pump of a secondembodiment of the present invention;

FIG. 15 is a perspective view showing a cross section taken along lineXV—XV in FIG. 14;

FIG. 16 is a graph of Pd(pressure) vs. h(specific enthalpy) forrefrigerant used in the system of the present invention;

FIG. 17 is another graph of Pd vs. h;

FIG. 18 is a schematic diagram showing a heat pump of a modified exampleof the second embodiment of the present invention;

FIG. 19 is a schematic diagram showing a heat pump of a modified exampleof the second embodiment of the present invention;

FIG. 20 is a schematic diagram showing a heat pump of a modified exampleof the second embodiment of the present invention;

FIG. 21 is a schematic diagram showing a heat pump of a third embodimentof the present invention;

FIG. 22 is a schematic diagram showing a heat pump of a fourthembodiment of the present invention;

FIG. 23 is a perspective view of a water heat exchanger of a sixthembodiment of the present invention;

FIG. 24 is a graph showing the relation between specific entropy andtemperature in the sixth embodiment;

FIG. 25 is a graph showing the relation between a temperature differenceΔT_(M) in the vicinity of an intermediate position in a water heatexchanger of the sixth embodiment and an optimum temperature differenceΔT₀;

FIG. 26 is a graph showing the relation between specific enthalpy andpressure in a seventh embodiment of the hot water supply systemaccording to the present invention;

FIG. 27 is a schematic diagram of the heat insulating tanks;

FIG. 28 is a schematic diagram showing the positions of temperaturesensors arranged in an exemplary heat insulating tank;

FIG. 29 is a schematic diagram showing the positions of temperaturesensors arranged in an exemplary heat insulating tank;

FIG. 30 is a schematic diagram showing the regulated condition of flowrate regulating valves for the heat insulating tanks;

FIG. 31 shows a modified example of a heat insulating section of thepresent invention;

FIG. 32 shows a modified example of a heat insulating section of thepresent invention; and

FIG. 33 is an alternate perspective view of a portion corresponding tothe portion shown in a cross-sectional view taken along line XV—XV inFIG. 14.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

(First Embodiment)

In the first embodiment of the present invention, a hot water supplysystem 100 for household use is shown in FIGS. 1 and 2. In FIG. 2,reference numeral 200 denotes a supercritical heat pump cycle 11 (“heatpump”) adapted to heat water to a high temperature (about 85° C.) foruse as a hot water generating fluid. The supercritical heat pump cycleis a heat pump cycle in which a refrigerant pressure on thehigh-pressure side becomes greater than or equal to a critical pressureof the refrigerant, and in which, for example, carbon dioxide, ethylene,ethane or nitrogen oxide is used as a refrigerant. Reference numeral 300denotes a plurality of heat insulating tanks for storing the waterheated by the heat pump 200. The heat insulating tanks 300 are arrangedto extend in parallel with a flow of hot water.

Referring to. FIG. 2, the heat pump 200 will now be described in moredetail. A compressor 210 for sucking and compressing a refrigerant(carbon dioxide in the present embodiment) is an electric compressorcomprising a unitary combination of a compression mechanism (not shown)for sucking and compressing the refrigerant, and an electric motor (notshown) for driving the compression mechanism. A water heat exchanger(radiator) 220 heat exchanges the refrigerant flowing out from thecompressor 210 and supplied water. The water heat ex changer 220 is acounterflow type heat exchanger formed so that a refrigerant flow isopposite that of a flow of water.

An electric expansion valve 230 decompresses the refrigerant flowing outfrom the water heat exchanger 220, and an evaporator 240 evaporates therefrigerant flowing out from the electric expansion valve 230 so thatthe refrigerant can absorb atmospheric heat before it is sent to anaccumulator 250 (suction side of the compressor 210) to be describedlater. An accumulator separates the refrigerant, which flows out fromthe evaporator 240, into a gas phase refrigerant and a liquid phaserefrigerant, before sending the gas phase refrigerant to the suctionside of the compressor 210 and storing excess refrigerant in the heatpump 200.

A blower 260 blows ambient air to the evaporator 240, and is capable ofregulating an air flow rate. The blower 260, the compressor 210 and theexpansion valve 230 are controlled by an electronic control unit (ECU)270 based on detected signals from various sensors which will bedescribed later. A refrigerant temperature sensor 271 detects thetemperature of the refrigerant flowing from the water heat exchanger220. A first water temperature sensor 272 detects the temperature of thewater flowing into the water heat exchanger.

A refrigerant pressure sensor 273 detects the pressure of therefrigerant flowing out of the water heat exchanger 220, and a secondwater temperature sensor 274 detects the temperature of the hot waterflowing from the water heat exchanger 220. The detected signals from thesensors 271-274 are input into the ECU 270.

The refrigerant pressure on the high-pressure side referred to above isa refrigerant pressure existing in a refrigerant passage extending froma discharge side of the compressor 210 to an inflow side of theexpansion valve 230, and is substantially equal to a discharge pressure(inner pressure of the water heat exchanger 220) of the compressor 210.The refrigerant pressure on the low-pressure side is present in arefrigerant passage extending from a discharge side of the expansionvalve 230 to a suction side of the compressor 210, and is substantiallyequal to a suction pressure of the compressor 210.

An electric water pump 400 supplies a regulated amount of hot water tothe water heat exchanger 220, and a shutoff valve 410 prevents tap watersupplied from a water pipe (not shown) from flowing into the water heatexchanger 220. Both the pump 400 and shutoff valve 410 are controlled bythe ECU 270.

Referring now to FIG. 3, the heat insulating tanks 300 will now bedescribed. The tank shown in FIG. 3 has a double tank structurecomprising outer and inner tank members 310, 320 made from a highcorrosion resistance metal, such as stainless steel, and a hollow space330 between the inner tank member 320 in which hot water is stored andthe outer tank member 310 is kept substantially vacuous.

The heat insulating tank 300 is provided at a lower portion thereof witha first opening 340 into which tap water flows, or from which the coldwater in the insulating tank 300 flows out toward the water heatexchanger 220. The heat insulating tank 300 is provided at an upperportion thereof with a second opening 350 from which the hot water inthe heat insulating tank 300 flows, or into which the water heated inthe water heat exchanger 220 flows.

The second opening 350 is provided with a heat insulating portion 360for preventing the heat of the hot water stored in the heat insulatingtank 300 from being discharged from the second opening 350 to theexterior (atmosphere). The heat-insulating portion 360 has a first pipesection 361 extending through the second opening 350 and into theinterior of the inner tank member 320, and a second pipe section 362extending from the first pipe section 361 to a pipe (not shown) joinedto the second opening 350. The two pipe sections 361, 362 are formed toan integral body by a low heat transfer rate resin.

At least one of the heat insulating tanks 300 is provided with aplurality of temperature sensors 370, which detect the temperature ofthe hot water therein, so that they are vertically spaced from oneanother. The temperatures detected by these temperature sensors 370 arealso input into the ECU 270.

The operation of the hot water supply system 100 of this embodiment willnow be described.

1. Hot Water Supplying Operation

1-1—Hot water stored in the heat insulating tanks is supplied:

Referring to FIG. 4A, when a hot water supply faucet (not shown) joinedto the side of the hot water supply system 100 (second openings 350 ofthe heat insulating tanks 300) is opened, the shutoff valve 410 isclosed, and tap water is supplied to the heat insulating tanks 300.

Consequently, the hot water stored in the heat insulating tanks 300 isforced out by the supplied tap water, and the hot water thus forced outis supplied from a hot water supply faucet.

1-2—Hot water temperature in the heat insulating tanks 300 decreases(the hot water is reheated):

As shown in FIG. 4B, when the temperature of the hot water in the heatinsulating tanks 300 is determined to be less than or equal to apredetermined level with reference to detected signals from thetemperature sensors 370, or when the volume of hot water is determinedto be less than or equal to a predetermined quantity, the shutoff valve410 is opened to operate the pump 400 and heat pump 200. Consequently,the hot water heated by the heat pump 200 flows into the heat insulatingtanks 300.

1-3—The hot water is re-heated as hot water is supplied:

As shown in FIG. 4C, the hot water supply faucet is opened to supply tapwater to the heat insulating tanks 300, while the shutoff valve 410 isopened to operate the pump 400 and heat pump 200. Consequently, the hotwater in the heat insulating tanks 300 and the water heated by the heatpump 200 are supplied.

2. Operation of the Heat Pump 200

When the compressor 210 is operated, the refrigerant is circulated inthe heat pump 200. Since the refrigerant discharged from the compressor210 at this time has been pressurized to a pressure greater than orequal to a critical level, it is circulated in the water heat exchanger220 without being condensed with a temperature gradient that is loweredgradually from a refrigerant inlet toward a refrigerant outlet.

Since the water heat exchanger 220 is formed so that a refrigerant flowand a flow of water to be supplied are oppositely directed, the water tobe supplied is heated in the water heat exchanger 220, with itstemperature increasing gradually from a water inlet toward a wateroutlet.

The refrigerant decompressed by the expansion valve 230 absorbs heatfrom the atmosphere and is evaporated in the evaporator 240, and theevaporated refrigerant is then sucked into the compressor 210 throughthe accumulator 250. In this embodiment, hot water of a temperaturegreater than or equal to 85° C. is supplied by the hot water supplysystem 100 (heat pump 200). Therefore, it is necessary during operationof the heat pump 200 that the temperature of the refrigerant at therefrigerant inlet of the water heat exchanger 220 be set to greater thanor equal to 85° C. In the present embodiment, the required refrigeranttemperature is preferably set to about 100° C. The temperature of thehot water to be supplied at the hot water outlet of the water heatexchanger 220 can be controlled by the pump 400 to a set level byregulating the quantity of the hot water to be supplied.

In the present embodiment, the degree of opening of the expansion valve230, rotational frequency of the compressor 210 and flow rate of theblower 260 are controlled so that a discharge pressure Pd (refrigerantpressure on the high-pressure side) of the compressor 210 reaches arefrigerant pressure corresponding to the required refrigeranttemperature at the refrigerant inlet of the water heat exchanger 220.

When the discharge pressure Pd is lower than a predetermined level Po(about 15 MPa in the present embodiment), it is controlled so that atemperature difference ΔT between the refrigerant flowing out from thewater heat exchanger 220 and water flowing thereinto reaches a level ina predetermined temperature range centering around a predeterminedtemperature difference ΔT₀ (about 10° in this embodiment).

When the temperature difference ΔT is larger than the predeterminedtemperature difference ΔT₀, the degree of opening of the expansion valve230 is reduced. The rotational frequency of the compressor 210 andairflow rate of the blower 260 may further be increased. The temperaturedifference ΔT is reduced by increasing the discharge pressure Pd in thismanner.

When the degree of opening of the expansion valve 230 is reduced, a flowpassage resistance of the refrigerant increases, so that the pressure ofthe refrigerant on the high-pressure side discharged from the compressor210 increases. When the airflow rate of the blower 260 is increased, thetransmission of heat from the outside air to the refrigerant isenhanced, so that the refrigerant pressure on the low-pressure sideincreases. Consequently, a suction rate of the refrigerant per unit timeof the compressor 210 increases to cause the refrigerant pressure on thehigh-pressure side to increase. Increasing the rotational frequency ofthe compressor 210 is also effective in increasing the refrigerantpressure on the high-pressure side. However, when the refrigerantpressure on the high-pressure side is increased by increasing therotational frequency of the compressor 210, the airflow rate of theblower 260 is also preferably simultaneously increased.

When the rotational frequency of the compressor 210 is merely increasedwithout varying the airflow rate thereof, a suction rate of therefrigerant increases but the refrigerant pressure on the low-pressureside decreases in consequence. Therefore, a mere increase in therotational frequency of the compressor 210 only minimally increases therefrigerant pressure on the high-pressure side. Increasing therefrigerant pressure on the low-pressure side by increasing the airflowrate of the blower 260 enables the compressor 210 to suck therefrigerant thereinto sufficiently to increase the refrigerant pressureon the high-pressure side.

When the discharge pressure Pd has reached a predetermined level Po, therefrigerant pressure on the low-pressure side is controlled so that thetemperature difference ΔT reaches a level in a predetermined rangearound the predetermined temperature difference ΔT₀. Specifically, theairflow rate of the blower is reduced, and the rotational frequency ofthe compressor 210 may further be increased. Thus, the refrigerantpressure on the low-pressure side is reduced.

When the airflow rate of the blower 260 is reduced, the quantity of heattransmitted from the outside air to the refrigerant decreases incontrast with the above-described example, so that the refrigerantpressure in the evaporator, i.e. the refrigerant pressure on thelow-pressure side, decreases. Increasing the rotational frequency of thecompressor 210 is also effective as mentioned above in reducing therefrigerant pressure on the low-pressure side.

The characteristics of the present embodiment will now be described.First, in order to have the following description understood easily, theheat exchange rate η of a counterflow type heat exchanger will bedescribed.

FIG. 5A is a schematic diagram of a counterflow type heat exchanger A,in which a heat exchanger fluid and a fluid to be heat-exchanged haveoppositely-directed flowpaths. If high-temperature water a andlow-temperature water b flow in outer and inner cylinders a, brespectively, the respective temperatures of the water a, b vary with asubstantially equal difference ΔT maintained as shown in FIG. 5B. Sincethe lateral axis x of FIG. 5B represents a distance based on theposition of a water inlet of the inner cylinder b, the quantity of heatQ transferred from the water a to the water; b in the heat exchanger Avaries in proportion to the area of the triangular hatched portion shownin FIG. 5B.

A quotient obtained by dividing a quantity of heat Q transferred to thefluid to be heat-exchanged (water b) by an average temperaturedifference At between the heat exchanger fluid (water a) and the fluidto be heat-exchanged is defined as heat exchange efficiency η(Q/Δt). Asis clear from this definition, a heat exchanger of a larger heatexchange efficiency η is capable of obtaining a larger heat exchangeamount with a smaller temperature difference.

A heat exchange efficiency of η the water heat exchanger 220 will now bediscussed in accordance with the above definition of heat exchangeefficiency η.

In the heat pump 200, the refrigerant pressure on the high-pressure sideis greater than or equal to a critical level. Accordingly, in the waterheat exchanger 220, the refrigerant varies (decreases) in temperaturealone without encountering the phase variation thereof.

Therefore, the temperature of the refrigerant varies substantiallylinearly with respect to the quantity of heat given to or removed fromthe refrigerant, in the same manner as when the refrigerant is in a gasphase or a liquid phase. Namely, assuming that the specific entropy ofthe refrigerant and refrigerant temperature vary substantially linearly,the temperature variations of the refrigerant and hot water becomesubstantially identical to the parameters shown in FIG. 5B.

A case where the heat pump 200 filled with a refrigerant comprisingfreon used at a pressure lower than a critical pressure is operated willnow be studied.

The freon varies in phase (condenses) at a constant temperature in atwo-phase region, i.e., in a gas-liquid phase region, and the specificentropy and temperature thereof vary substantially linearly in a heatedvapor region and an over-cooled region.

When a temperature difference between the freon and hot water in aheated vapor region, i.e. a temperature difference between thetemperature of hot water at the hot water outlet side of the water heatexchanger 220 and the temperature of the freon at the refrigerant inletside thereof is set to Δt₁, the hot water temperature variesexponentially in the gas-liquid phase region as shown in FIG. 6.Therefore, in order to conduct heat exchange reliably between the freonand hot water in the water heat exchanger 220, it is necessary that atemperature difference Δt₂ between the water inlet side and refrigerantinlet side be set sufficiently high as compared with the temperaturedifference Δt₁.

Therefore, when a heat pump 200 containing a refrigerant comprisingfreon used at a pressure lower than a critical pressure is operated, theheat exchange efficiency η decreases as compared with that in thepresent embodiment in which the heat pump 200 filled with a refrigerant,such as carbon dioxide, used at a pressure greater than or equal to acritical level, is operated.

In the hot water supply system 100 in this embodiment, water is heatedwith the heat absorbed from the atmosphere, and by the supercriticalheat pump cycle 200 of a high heat exchange efficiency η, so that theamount of electric power required to heat water can be minimized.

Moreover, since hot water can be generated even in the daytime duringpeak power fee rates, it is unnecessary to store hot water for daytimeuse in large heat insulating tanks. Accordingly, the hot water supplysystem 100 can be installed in a smaller spaces than conventionalheating systems.

Referring to FIG. 7, although the above description is given based onthe assumption that the specific entropy and temperature of therefrigerant vary linearly, the specific entropy and the refrigeranttemperature practically varies so as to increase non-linearly inaccordance with an increase in the specific entropy thereof. As is clearfrom FIG. 7, the graph showing the relation between the specific entropyof the refrigerant and the temperature thereof indicates that theproperty of the refrigerant varies in accordance with the pressurethereof, and that, as the refrigerant pressure increases, the relationbetween the specific entropy of the refrigerant and the temperaturethereof comes close to a linearly varying condition.

As is clear from the above description, the heat exchange efficiency ηcan be increased when the specific entropy of the refrigerant andtemperature thereof vary linearly, so that, when the refrigerantpressure, i.e. the discharge pressure Pd, is increased, the heatexchange efficiency η can be increased.

Therefore, in this embodiment, the discharge pressure Pd is controlledso that the temperature difference ΔT between the refrigerant flowingfrom the water heat exchanger 220 and the hot water flowing thereintoreaches a predetermined level ΔT₀ to improve the heat exchangeefficiency η. The reasons why such techniques are used will now begiven.

The solid lines a-b-c-d in. FIG. 7 form a heat pump cycle diagram duringa period of time in which the heat pump 200 is stable in a certaincondition (discharge pressure Pd=9 Mpa).

Referring to the chart, line segment a-b indicates a compression stroke,line segment b-c a change in the interior of the water heat exchanger,line segment c-d a decompression stroke, line segment d-a an evaporationstroke, and broken line A-B a change of hot water in the water heatexchanger.

When the discharge pressure Pd is increased from 9 MPa to 11 MPa, theheat exchange efficiency η is improved to cause a heat exchange rate inthe water heat exchanger 220 to increase. Consequently, the temperatureand specific entropy of the refrigerant on the refrigerant outlet sideof the water -heat exchanger 220 decrease as shown in FIG. 8.

Namely, as is clear from FIGS. 7 and 8, when the heat exchangeefficiency η is improved to cause the heat exchange rate in the waterheat exchanger 220 to increase, a temperature difference ΔT between therefrigerant flowing out from the water heat exchanger 220 and waterflowing thereinto becomes small. Therefore, in this embodiment, the heatexchange efficiency η is improved by controlling the discharge pressurePd so that the temperature difference ΔT attains a predetermined levelΔT₀, whereby the efficiency of the heat pump 200 (hot water supplysystem 100) is improved. When the temperature difference A T is small,the heat exchange rate in the water heat exchanger 220 increases, sothat the rotational frequency (the feed rate of hot water to besupplied) of the pump 400 can be increased.

When the refrigerant pressure on the low-pressure side of the heat pump200 increases due to, for example, an increase in the temperature of theoutside air when the heat pump 200 is operated in the stable conditionshown in FIG. 7, the temperature of the refrigerant discharged from thecompressor 210 decreases as shown in FIG. 9 when the discharge pressurePd is identical.

When the discharge pressure Pd is increased at this time to increase thetemperature of the refrigerant discharged from the compressor 210, thedischarge pressure Pd increases excessively as shown in FIG. 10.Therefore, this excessive pressure may overcome the pressure resistingstrength of the compressor 210 and water heat exchanger 220. Althoughthe pressure resisting strength of the compressor 210 and water heatexchanger 220 may be increased to solve this problem, the cost ofmanufacturing the hot water supply system 100 would consequentlyincrease.

In the present embodiment, when the discharge pressure Pd is lower thana predetermined level P₀, the pressure Pd is controlled so that thetemperature difference ΔT attains a predetermined level ΔT₀. When thedischarge pressure Pd has reached the predetermined level P₀, therefrigerant pressure on the low-pressure side is controlled so that thetemperature difference ΔT reaches the predetermined level ΔT₀, with therefrigerant temperature maintained at a level greater than or equal to arequired level. As a result, an excessive increase in the dischargepressure Pd can be prevented as shown in FIG. 11. Namely, when therefrigerant pressure on the low-pressure side is reduced, therefrigerant temperature can be increased to a level greater than orequal to a required level without causing the refrigerant pressure onthe high-pressure side to increase excessively.

Therefore, since it is unnecessary in this embodiment to increase thepressure resisting strength of the compressor 210 and water heatexchanger 220, the hot water supply system 100 (heat pump 200) can beoperated efficiently while minimizing an increase in the manufacturingcost.

As is clear from the above description, the predetermined temperaturedifference ΔT₀ does not have a fixed value, but rather varies dependingupon the heating capacity and the refrigerant pressure on thehigh-pressure side, so that it may be varied in accordance with theseconditions.

Since the second openings 350, into and out of which high-temperaturehot water flows, of the heat insulating tanks 300 are covered with theheat insulating portions 360 comprising a resin material of a low heatconductivity, the radiation of the heat of the hot water stored in theheat insulating tanks 300 from the second openings 350 to the exterior(into the atmosphere) can be minimized. This enables the heat insulatingcapacity of the heat insulating tanks 300 to be improved.

As the temperature sensors 370 for detecting the temperature of the hotwater in the heat insulating tanks 300 are be vertically spaced from oneanother, the temperature and volume of the hot water in the heatinsulating tanks 300 can be accurately determined. Therefore, operationof the heat pump 200 can be accurately controlled.

In this embodiment, the heat insulating tanks 300 are arranged inparallel with the flow of hot water, and the condition (temperature andvolume of hot water) of each heat insulating tank 300 can be regarded asbeing substantially identical. Thus, the temperature sensors 370 are notprovided in each of the plural heat insulating tanks 300, but rather areprovided in one heat insulating tank 300. Therefore, the heat pump 200is controlled based on the condition of the heat insulating tank 300provided with the temperature sensors 370 therein.

In this embodiment, the rotational frequency of the pump 400 iscontrolled based on the temperature detected by the temperature sensor274. Also, the rotational frequency (supply rate of hot water to thewater heat exchanger 220) of the pump 400 may be set to a constant ratewithout the need to rely upon the operational condition of the heat pump200.

When the temperature difference ΔT immediately after starting of theheat pump 200 is large, the discharge pressure Pd as well as thetemperature of the hot water flowing out from the water heat exchangeris low as shown in FIG. 12. However, when the expansion valve 230 andcompressor 210 are controlled so that the temperature difference ΔTbecomes less than or equal to the predetermined level ΔT₀, therefrigerant pressure on the high-pressure side increases as shown inFIG. 13. Accordingly, the efficiency of the heat pump 200 can beimproved by increasing the heat exchange efficiency η while obtaininghigh-temperature hot water.

It is also allowable to fix the rotational frequency of the pump 400,and to provide a throttle, the degree of opening of which isregulatable, in a pump flow passage to control the flow rate of the hotwater.

(Second Embodiment)

In the above-described first embodiment, the refrigerant pressure on thelow-pressure side of the heat pump 200 is reduced when it increases dueto an increase in the temperature of the outside air, and the heat pump200 is controlled so that the discharge pressure Pd does not exceed thepredetermined level P₀. In the second embodiment, heat exchange isconducted between the refrigerant between the refrigerant inlet andoutlet of a water heat exchanger 220 and the refrigerant sucked into acompressor 210.

FIG. 14 is a schematic diagram showing a heat pump 200 alone in a hotwater supply system 100 of this embodiment. A refrigerant heat exchanger280 conducts heat exchange between the refrigerant existing between therefrigerant inlet and outlet of the water heat exchanger 220 and therefrigerant sucked into the compressor 210.

FIG. 15 is a perspective view showing a section taken along line XV—XVin FIG. 14. As shown, the water heat exchanger 220 comprises arefrigerant side multi-hole-carrying tube 222 provided with a pluralityof holes 221 through which the refrigerant is circulated, and a hotwater side multi-hole-carrying tube 224 provided with a plurality ofholes 223 through which the hot water flows. The refrigerant heatexchanger 280 has the same structure as the refrigerant sidemulti-hole-carrying tube 222, and is joined thereto. The twomulti-hole-carrying tubes 222, 224 and refrigerant heat exchanger 280are molded by subjecting aluminum to an extrusion or drawing process.

The characteristics of this embodiment will now be described. When thetemperature of the outside air is high during summer, the refrigerantpressure on the low-pressure side of the heat pump 200 increases.Therefore, to obtain hot water of a temperature of greater than or equalto the predetermined level (85° C.), it is necessary that the dischargepressure Pd be increased as well in accordance with an increase in therefrigerant pressure on the low-pressure side as shown by a-b-c-d inFIG. 16.

When heat exchange is conducted as mentioned above between therefrigerant sucked into the compressor 210 and the refrigerant existingbetween, the refrigerant inlet and outlet of the water heat exchanger220, the temperature of the suction refrigerant increases as shown byA-B-C-D in FIG. 16, so that the refrigerant temperature on the dischargeside of the compressor 210 can be increased without causing thedischarge pressure Pd to increase. Accordingly, it is unnecessary toincrease the pressure resisting:strength of the compressor 210 and waterheat exchanger 220, so that an increase in the cost of manufacturing thehot water supply system 100 can be minimized.

In this embodiment, the discharge temperature is increased, as anincrease in the discharge pressure Pd is suppressed, by increasing thetemperature of the suction refrigerant. Therefore, increasing thetemperature of the suction refrigerant by subjecting the suctionrefrigerant and the refrigerant from the water heat exchanger 220 toheat exchange is conceivable.

However, in the heat pump 200, the refrigerant in the water heatexchanger 220 is circulated so as to have a temperature gradient inwhich the temperature thereof decreases gradually from the refrigerantinlet toward the refrigerant outlet as mentioned above, so that atemperature difference between the discharge refrigerant and the hotwater flowing into the water heat exchanger 220 is small.

Since the temperature of the outside air is high, a temperaturedifference between the refrigerant (suction refrigerant) in anevaporator 240 and the outside air is also small.

Therefore, the device for conducting heat exchange between the dischargerefrigerant and suction refrigerant is not, capable of sufficientlyincreasing the temperature of the suction refrigerant, thus possiblycausing the discharge pressure Pd to increase excessively.

Increasing the temperature of the suction refrigerant by subjecting therefrigerant which is going to flow into the water heat exchanger 220 andthe suction refrigerant is conceivable but, when such a device is used,the temperature of the refrigerant flowing into the water heat exchanger220 decreases, and hot water of a predetermined temperature cannot beobtained.

The present embodiment can be utilized not only in summer during whichthe temperature of the outside air is high but also in spring, autumnand winter in which the temperature of the outside air is not so high(refer to FIG. 17).

In this embodiment, the suction refrigerant is heated with therefrigerant circulated through the water heat exchanger 220 by bringingthe refrigerant heat exchanger 280 into direct contact with therefrigerant side of the multi-hole-carrying tube 222. As shown in FIG.18, this embodiment may also be formed so that the refrigerantcirculated through the water heat exchanger 220 is taken out therefromduring circulation thereof and subjected to heat exchange with thesuction refrigerant.

As shown in FIGS. 19 and 20, a bypass passage 290 for introducing therefrigerant, which flows out from the accumulator 250, into the suctionside of the compressor 210 via a long way around the refrigerant heatexchanger 280, and an electromagnetic valve 291 for closing this bypasspassage 290 may be provided.

When the temperature of the outside air is high in the summer, thesuction refrigerant is heated with the electromagnetic valve 291 closed.When the temperature of the outside air is not so high in the spring,autumn and winter, it is preferable to stop heating of the suctionrefrigerant with the electromagnetic valve 291 opened.

FIG. 19 shows an example in which a bypass passage 290 is provided inthe same heat pump 200 as shown in FIG. 14, and FIG. 20 shows an examplein which a bypass 290 is provided in the same heat pump 200 shown inFIG. 18.

In the heat pump 200 shown in FIG. 18, a bypass passage 290 and anelectromagnetic valve 291 may be provided in the refrigerant passageextending from the water heat exchanger 220 to the refrigerant heatexchanger 280 to control an operation for determining whether thesuction refrigerant should be heated.

(Third Embodiment)

In the second embodiment, heat exchange is conducted in the refrigerantheat exchanger 280 between the refrigerant between the refrigerant inletand outlet of the water heat exchanger and the refrigerant; sucked intothe compressor 210. In the third embodiment, a heat exchange rate in arefrigerant heat exchanger 280 is controlled based on the condition ofthe refrigerant on the discharge side of a compressor 210.

Referring to FIG. 21, in this embodiment, a refrigerant passage 281 forintroducing a liquid-phase refrigerant collected in a bottom portion ofan accumulator 250 into the refrigerant flowing out thereof, and anelectric flow rate regulating valve 282 for regulating the communicationcondition of the refrigerant passage 281 are provided.

The compressor 210 is provided on the discharge side thereof with atemperature sensor 283 for detecting the temperature of the dischargerefrigerant, and a detected temperature is input into an electroniccontrol unit (ECU) 285 adapted to control the flow rate regulating valve282.

The operation of this embodiment will now be described. In the secondembodiment, heat exchange is conducted in the refrigerant heat exchanger280 between the refrigerant between the refrigerant inlet and outlet ofthe water heat exchanger 220 and that sucked into the compressor 210, sothat a heat exchange rate is determined based on the capacity of therefrigerant heat exchanger 280.

Therefore, if the capacity of the refrigerant heat exchanger 280 isdetermined when summer temperatures cause the ambient air to be high,the degree to which the suction refrigerant is heated during lowtemperature becomes high (refer to FIGS. 16 and 17) as compared to aperiod of time in which the temperature of, the outside air is high, sothat the temperature of the discharged refrigerant becomes higher thannecessary.

When a temperature detected by a temperature sensor 283 has becomehigher than a predetermined level (about 100° C. when hot water of 85°C. is obtained), the degree of opening of the flow rate regulating valve282 is increased to increase the supply rate of the liquid-phaserefrigerant sent from the accumulator 250 to the refrigerant heatexchanger 280, to thereby prevent the degree of heating of the suctionrefrigerant from becoming higher than necessary.

When the temperature detected by the temperature sensor 283 is less thanor equal to a predetermined level, the degree of opening of the flowrate regulating valve 282 is reduced to lower the supply rate of theliquid-phase refrigerant sent from the accumulator 250 to therefrigerant heat exchanger 280, and to thereby prevent the degree ofheating the suction refrigerant from becoming lower than necessary.

Since the above operation enables the discharge temperature to bemaintained at a predetermined level independent of the temperature ofthe outside air, hot water of a predetermined temperature can besupplied in a stable manner as an excessive increase in the dischargepressure Pd is prevented.

(Fourth Embodiment)

In the third embodiment, an increase in the temperature of the dischargerefrigerant to a level higher than necessary is electronically preventedby the temperature sensor 283 and flow rate regulating valve 282.Referring specifically to FIG. 22, the fourth embodiment mechanicallyprevents the temperature of the discharge refrigerant from increasing toa level higher than necessary based on the fact that isoentropic curvesbecome substantially parallel on a gas-phase region side of a saturatedair line.

This embodiment has a flow rate regulating valve 282 with a temperaturesensing cylinder 282 a such as a known temperature-type expansion valveused in a heat pump cycle using freon as a refrigerant. The supply rateof the liquid-phase refrigerant sent from an accumulator 250 to therefrigerant heat exchanger 280 is regulated by the flow rate regulatingvalve 282 so that the degree of heating between an outlet side of arefrigerant heat exchanger 280 and an inlet side of a compressor 210reaches a predetermined level.

The structure of the fourth embodiment thus enables system manufacturingcosts to be reduced, and enables hot water of a predeterminedtemperature to be supplied in a stable manner while preventing thedischarge pressure Pd from increasing excessively.

(Fifth Embodiment)

In the first embodiment, the temperature difference ΔT between therefrigerant flowing from the water heat exchanger 220 and the waterflowing thereinto is controlled by varying the degree of opening of theexpansion valve 230. Also, the temperature of hot water at the hot wateroutlet of the water heat exchanger 220 is controlled by regulating theflow rate of the hot water by the pump 400.

In the fifth embodiment, a temperature difference ΔT between arefrigerant flowing from a water heat exchanger 220 and the waterflowing thereinto is controlled to a predetermined level ΔT₀ byregulating a flow rate of the hot water by a pump 400. The temperatureof hot water is controlled by regulating the degree of opening of anexpansion valve 230.

The methodology associated with the present embodiment will now bedescribed.

First, the temperature difference ΔT between the refrigerant flowing outof the water heat exchanger 220 and the water flowing thereinto isdetected, and it is determined whether this temperature difference ΔT isin a predetermined temperature range (ΔT₀±α) centering around apredetermined temperature difference ΔT₀. When the temperaturedifference ΔT is in the predetermined temperature range, the flow rateof hot water regulated by the pump 400 is kept constant. When it is notin the predetermined temperature range, i.e., when the temperature ofthe refrigerant flowing out of the water heat exchanger 220 is higherthan that of the water flowing thereinto by a level greater than thepredetermined temperature difference ΔT₀, the flow rate of the hot wateris increased by the pump 400 so as to lower the temperature of therefrigerant. As a result, the transfer of heat from the refrigerant tothe hot water is promoted in the water heat exchanger 220, and thetemperature of the refrigerant flowing out of the water heat exchanger220 decreases. When the temperature difference ΔT between therefrigerant and the water enters the temperature range centering aroundthe predetermined temperature difference ΔT₀ due to the decrease in thetemperature of the refrigerant, the pump 400 maintains the flow rate ofthe hot water at that time.

The temperature Twh of the hot water at the hot water outlet of thewater heat exchanger 220 is then detected, and it is determined whetherthis temperature Twh corresponds to a set temperature Tw₀. When thetemperature Twh of the hot water corresponds to the set temperature Tw₀,the degree of opening of the expansion valve 230 is retained. When thetemperature Twh of the hot water is not higher than the set temperatureTw₀, the discharge pressure Pd of the compressor 210 is increased byreducing the degree of opening of the expansion valve 230, to increasethe temperature of the refrigerant flowing into the water heat exchanger220. When the discharge pressure Pd reaches a predetermined level P₀(for example, 15 MPa) during this time, the refrigerant pressure on thelow-pressure side is reduced by increasing the rotational frequency ofthe compressor 210 and/or reducing the airflow rate of a blower 260.This enables the temperature of the refrigerant discharged from thecompressor 210 to increase without causing the refrigerant pressure onthe high-pressure side to excessively increase.

When the temperature difference ΔT between the refrigerant and the waterdoes not fall within the temperature range centering around thepredetermined temperature difference ΔT₀ after such a refrigerantpressure control operation, the above-mentioned flow rate control of thepump 400 is again performed. Finally, the hot water supply system 100 isput in the condition in which the temperature difference ΔT between therefrigerant and the water falls within a temperature range centeringaround the predetermined temperature difference ΔT₀ with the temperatureTwh of hot water supplied corresponding to the set level Two.

In this embodiment, the pressure Pd of the refrigerant discharged fromthe compressor 210 is controlled by regulating the degree of opening ofthe expansion valve 230 so that the temperature Twh of the hot watersupplied reaches a set level Two. However, when the temperaturedifference ΔT between the refrigerant and the water is larger than the-predetermined temperature difference ΔT₀, the flow rate of the hot wateris increased by operating the pump 400. As a result, the temperature Twhof the hot water also decreases to finally increase the refrigerantpressure on the high pressure side or decrease the refrigerant pressureon the low-pressure side.

Therefore the control operation in this embodiment can be called acontrol operation in which at least one of the refrigerant pressure onthe high-pressure side and that on the low-pressure side is controlledso that the temperature difference ΔT between the refrigerant flowingout of the water heat exchanger 220 and the water flowing thereintoattains a level in a range including the predetermined level ΔT₀.

(Sixth Embodiment)

In the first embodiment, the degree of opening of the expansion valve230 is controlled so that the temperature difference ΔT between therefrigerant flowing out of the water heat exchanger 220 and the waterflowing thereinto reaches the predetermined level ΔT₀.

As described in connection with the first embodiment, it is necessarythat the temperature difference ΔT between the refrigerant flowing fromthe water heat exchanger 220 and the water flowing thereinto be set tothe predetermined level ΔT₀ to improve the heat exchange efficiency ofthe water heat exchanger 220. An optimum level of the temperaturedifference ΔT between the refrigerant and water varies depending uponthe heating capacity of the heat pump cycle and the flow rate of the hotwater.

In the sixth embodiment, an optimum temperature difference ΔT_(A)between the refrigerant and water is computed, and a temperaturedifference ΔT between the refrigerant flowing from a water heatexchanger 220 and the hot water flowing thereinto is controlled so thatit reaches a level within a predetermined temperature range centeringaround this optimum temperature difference ΔT_(A).

A method of computing the optimum temperature difference ΔT_(A) will nowbe described.

As shown in FIG. 23, the water heat exchanger 220 in the sixthembodiment is formed by pasting together a hot water sidemulti-hole-carrying plate 501 provided with a plurality of holes throughwhich hot water flows, and a refrigerant side multi-hole-carrying plate502 provided with a plurality of holes through which a refrigerant iscirculated. The hot water side and refrigerant side multi-hole-carryingplates 501, 502 are provided at substantially intermediate portionsthereof in the flow directions of hot water and refrigerant withtemperature sensors 511, 512, which are adapted to detect thetemperatures of the hot water and refrigerant which flow in theintermediate portion of the water heat exchanger 220. Accordingly, atemperature difference ΔT_(M) between the hot water and refrigerant inthe intermediate portion of the water heat exchanger 220 can bedetermined by computing a difference between the hot water temperatureand refrigerant temperature detected by the temperature sensors 511,512.

According to a temperature gradient shown in FIG. 24 in the refrigerantside multi-hole-carrying plate 502, the refrigerant temperature and hotwater temperature converge toward one another in the vicinity of theintermediate portion thereof. Specifically, the temperature differenceΔT_(M) between the refrigerant and hot water in a position in thevicinity of the intermediate portion of the water heat exchanger 220 isa value most clearly representing the heating capacity of the water heatexchanger 220.

Therefore, when a target value of the temperature difference ΔT₀ betweenthe refrigerant flowing from the water heat exchanger 220 and the waterflowing thereinto is set based on the temperature difference between thehot water and refrigerant in a position in the vicinity of theintermediate portion of the water heat exchanger, an optimum temperaturedifference ΔT_(A) can be set as a target value.

The temperature difference ΔT_(M) in a position in the vicinity of theintermediate portion of the water heat exchanger and optimum temperaturedifference ΔT_(A) have a correlation shown in FIG. 25, and the optimumtemperature difference ΔT_(A) is therefore determined based on adetected temperature difference ΔT_(M) with reference to the relationshown in FIG. 25. When the actually detected temperature difference ΔT₀is larger than the optimum temperature difference ΔT_(A), the degree ofopening of an expansion valve 230, airflow rate of a blower 260 androtational frequency of a compressor 210 are changed to increase therefrigerant pressure on the high-pressure side or reduce the refrigerantpressure on the low-pressure side. As a result, the heat exchangeefficiency in the water heat exchanger 220 increases, so that thetemperature difference ΔT₀ approaches the optimum temperature differenceΔT_(A). When the actually detected temperature difference ΔT₀ is smallerthan the optimum temperature difference ΔT_(A), an efficient operatingcondition is not obtained, and the operation of a heat pump cycle 200 ischanged by reducing the refrigerant pressure on the high-pressure sideor increasing the refrigerant pressure on the low-pressure side.

It is preferable that the optimum temperature difference ΔT_(A) be setwith consideration given not only to the temperature difference ΔT_(M)in the vicinity of the intermediate portion of the water heat exchanger220, but also to the temperature of the water flowing into the waterheat exchanger 220, temperature (temperature of hot water supplied) ofthe hot water flowing therefrom, and temperature of the refrigerant onthe low-pressure side. Therefore, the sixth embodiment is provided witha temperature sensor for detecting the temperature of the refrigerant onthe low-pressure side in addition to the structure identical with thatof the first embodiment.

There is a tendency for the optimum temperature difference ΔT_(A) tobecome large as, for example, the temperature of the water flowing intothe water heat exchanger 220 increases. The optimum temperaturedifference ΔT_(A) also tends to become large as the temperature of thehot water flowing out of the water heat exchanger 220 increases and asthe temperature of the refrigerant on the low-pressure side increases.Namely, when these conditions are satisfied, the heat pump cycle 200 isnecessarily operated in an inefficient region, so that the heat exchangeefficiency decreases to cause the optimum temperature difference ΔT_(A)to increase. Therefore, when these parameters are taken intoconsideration, the optimum temperature difference ΔT_(A) can beaccurately set.

(Seventh Embodiment)

In the first embodiment, the degree of opening of the expansion valve230 is controlled so that the temperature difference ΔT between therefrigerant flowing from the water heat exchanger 220 and the waterflowing thereinto reaches the predetermined level ΔT₀. Morespecifically, when the temperature difference ΔT is larger than thetemperature difference ΔT₀, the heat exchange efficiency of the waterheat exchanger 220 is increased by increasing the refrigerant pressure(discharge pressure Pd of the compressor. 210), whereby a controloperation is carried out to reduce the temperature difference ΔT to thelevel ΔT₀.

As shown by a broken line X in. FIG. 26, this temperature difference ΔTdecreases into a certain region when the refrigerant pressure on thehigh-pressure side is increased. However, it thereafter increases whenthe refrigerant pressure on the high-pressure side continues toincrease. The reasons why this phenomenon occurs are explained asfollows.

Since a heat transfer rate α in the water heat exchanger 220 is high inthe vicinity of saturation curves, the quantity of heat transferred fromthe refrigerant to hot water increases and the temperature difference ΔTdecreases as the refrigerant pressure on the high-pressure sideincreases. However, in a region away from the saturation curves, theheat transfer rate α decreases, so that, no matter how much therefrigerant pressure on the high-pressure side is increased, the heatthereof is not effectively transmitted to the hot water.

Therefore, in the seventh embodiment, an ECU 270 determines whether thetemperature difference ΔT becomes small/large with respect to adecrease/increase in the degree of opening of an expansion valve 230,and determines whether to reduce or increase the degree of opening ofthe expansion valve 230 in accordance with an actual temperaturedifference ΔT.

Namely, when the temperature difference ΔT is higher than thepredetermined level ΔT, when the temperature difference ΔT becomes largewith respect to a decrease in the degree of opening of the expansionvalve 230, or, when ΔT becomes small with respect to an increase in thedegree of opening of the expansion valve, the degree of opening thereofis increased. When the temperature difference ΔT is lower than thepredetermined level ΔT₀ in the same case, the degree of opening of theexpansion valve is reduced.

On the other hand, when the temperature difference ΔT is higher than thepredetermined level ΔT₀ when the temperature difference ΔT becomes smallwith respect to a decrease in the degree of opening of the expansionvalve 230, or, when the temperature difference ΔT becomes large withrespect to an increase in the degree of opening of the expansion valve230, the degree of opening thereof is reduced. When the temperaturedifference ΔT is lower than the predetermined level ΔT₀ in the samecase, the degree of opening of the expansion valve 230 is increased.

As a result of the above control,:the temperature difference ΔT betweenthe refrigerant flowing from the water heat exchanger 220 and waterflowing thereinto can always be controlled to the level ΔT₀.

(Other Embodiments)

In the above-described embodiments, the heat insulating tanks 300 areadjacent to the heat pump 200 as shown in FIG. 1. The heat insulatingtanks 300 may also be provided under a floor of a house as shown in FIG.27.

Also, in the above-described embodiments, the temperature sensors 370are provided in the interior of the inner tank member 320. However, thesensors may also be provided on an outer surface of the inner tankmember 320 as shown in FIG. 28.

As shown in FIG. 29, both first and second openings 340, 350 may beprovided concentrically at lower sides so that the temperature sensors370 may be fixed on a pipe 351 extending upwardly through the firstopening 340.

As shown in FIG. 30, flow rate regulating valves 380 may be provided onthe side of first openings 340 so that the flow rates of tap waterentering the heat insulating tanks 300 become substantially equal.

Such arrangements can prevent a large difference between conditions inthe heat insulating tank provided with the temperature sensors 370therein and those in heat insulating tanks not provided therewith, sothat accurate control of the heat pump 200 may be performed.

In the above-described embodiments, the heat insulating portion 360 isformed by a resin. The heat insulating portion 360 may also be formed byvertically bending a passage 351 in a second opening 350 into alabyrinth structure, as shown FIG. 31.

Since the passage 351 in this structure is vertically bent,high-temperature hot water is positioned in an upper section of aportion A of the passage 351, and low-temperature hot water in a lowersection thereof. Therefore, the length of the passage 351 increases,and, moreover, natural convection rarely occurs in the passage 351, thusenabling the heat insulating stress of the heat insulating tanks 3.00 tobe improved.

As shown in FIG. 32, a valve for closing the second opening 350 when hotwater is not circulated therethrough may be formed by a material withhigh heat insulating characteristics, such as a resin.

Referring to FIG. 32, a first valve disc 361 of a resin for closing thesecond opening 350 is provided with a through passage 362 extendingtherethrough.

A first coiled spring 363 applies a resilient force directed to closethe first opening 350 to an inner side of the relative heat insulatingtank 300. A second valve disc 364 of a resin is provided in the portionof the through passage 362 on an inner side of the heat insulating tank300, and closes or opens the through passage 362. A second coiled springapplies a resilient force directed to close the through passage 362 tothe second valve disc 364.

As a result of the above-described structure, the first valve disc 361is opened when the hot water flows out of the heat insulating tank 300,and the second valve disc 364 when the water heated by the heat pump 200flows thereinto. Both of the valve discs 361, 364 are closed when thehot water is not circulated.

Also, in the second embodiment, each of the water heat exchanger 220 andrefrigerant heat exchanger 280 may be a triple cylinder type heatexchanger (FIG. 33).

In the above-described embodiments, water is heated by the heat pump200. The use of the heat pump 200 is not limited to those described inthe above embodiments; i.e., the heat pump 200 may also be used to heatair and other types of fluids.

The above-described embodiments refer to examples in which the water isheated by the heat pump cycle 200 and stored in the heat insulatingtanks 300. However, the above-described heat pump cycles 200 can also beapplied to hot water supply systems which are other than theabove-mentioned types. For example, a system for heating a heataccumulating refrigerant as a hot water generating fluid by the heatpump cycle 200, storing the resultant heat accumulating refrigerant in aheat insulating tank or an open-atmosphere type hot water storage tank,and heating the water by conducting heat exchange between the same andthe resultant heat accumulating refrigerant may be utilized. In such anindirect heat exchange type hot water supply system, the hot water to besupplied does not require storage, so that this hot water supply systemis highly sanitary. Since, unlike tap water, heat accumulatingrefrigerant is used with no predetermined pressure applied thereto, thehot water storage tanks for storing the heat accumulating refrigerantneed not be high strength tanks such as those required for storingheated city tap water require. Therefore, the cost of manufacturing thehot water storage tanks can be minimized.

While the above description constitutes the preferred embodiment of thepresent invention, it should be appreciated that the invention may bemodified without departing from the proper scope or fair meaning of theaccompanying claims. Various other advantages of the present inventionwill become apparent to those skilled in the art after having thebenefit of studying the foregoing text and drawings in conjunction withthe following claims.

What is claimed is:
 1. A hot water supply system for heating hot watergenerating fluid via a heat pump cycle in which a refrigerant on ahigh-pressure side thereof reaches a pressure level greater than orequal to a critical pressure of the refrigerant, comprising: acompressor for sucking and compressing a refrigerant; a radiator forsubjecting the compressed refrigerant and the hot water generating fluidto heat exchange, and for defining flow paths so that thecompressor-discharged refrigerant and the hot water generating fluidflow in opposite directions; a decompressor for decompressing therefrigerant flowing from said radiator; an evaporator for evaporatingthe refrigerant flowing from said decompressor, to cause the refrigerantto absorb heat, and to output the refrigerant toward a suction side ofsaid compressor; and a controller for controlling the refrigerantpressure on the high-pressure side so that a temperature difference ΔTbetween the refrigerant flowing from said radiator and the hot watergenerating fluid flowing into said radiator reaches a predeterminedlevel ΔT₀.
 2. The system of claim 1, wherein when the refrigerantpressure on the high-pressure side is higher than a predetermined level,said controller controls refrigerant pressure on a low-pressure side sothat the temperature difference ΔT reaches the level ΔT₀.
 3. The systemof claim 1, further comprising a pump for controlling a flow of the hotwater generating fluid to the radiator, wherein the controller controlsrefrigerant pressure on a low-pressure side so that the temperaturedifference ΔT reaches the predetermined level ΔT₀ by varying operationof the decompressor, and controls a temperature of the hot watergenerating fluid at an outlet of the radiator by varying operation ofthe pump.
 4. The system of claim 1, further comprising a pump forcontrolling a flow of the hot water generating fluid to the radiator,wherein the controller controls the flow of the hot water generatingfluid so that the temperature difference ΔT reaches the predeterminedlevel ΔT₀, and controls a temperature of the hot water generating fluidat an outlet of the radiator by varying operation of the decompressor.5. The system of claim 1, wherein the controller controls at least oneof the refrigerant pressure on the high-pressure side and refrigerantpressure on a low-pressure side so that the temperature difference ΔTreaches the predetermined level ΔT₀.
 6. The system of claim 1, whereinthe controller controls a rotational frequency of the compressor tocontrol a discharge pressure of the refrigerant discharged from thecompressor.
 7. The system of claim 1, further comprising means forregulating a flow rate of air sent to said evaporator, whereinrefrigerant pressure on a low-pressure side is controlled by varying theflow rate of the air supplied to said evaporator via said flow rateregulating means.
 8. The system of claim 7, wherein said flow rateregulating means comprises a blower.
 9. The system of claim 1, wherein aflow rate of the refrigerant circulated through said radiator isregulated so that a temperature of the refrigerant flowing from saidradiator reaches a level greater than or equal to a predetermined level.10. The system of claim 1, wherein said controller regulates a degree ofopening- of said decompressor, and the refrigerant pressure on thehigh-pressure side is controlled by varying the degree of opening ofsaid decompressor.
 11. The hot water supply system according to claim10, wherein: when the temperature difference is increased by a decreasein the degree of opening of the decompressor or the temperaturedifference is decreased by an increase in the degree of opening of thedecompressor, the degree of opening of the decompressor is increasedwhen the temperature difference is larger than the predeterminedtemperature difference, and the degree of opening of the decompressor isdecreased when the temperature difference is smaller than thepredetermined temperature difference; and when the temperaturedifference is decreased by a decrease in the degree of opening of thedecompressor or the temperature difference is increased by an increasein the degree of opening of the decompressor, the degree of opening ofthe decompressor is decreased when the temperature difference is largerthan the predetermined temperature difference, and the degree of openingof the decompressor is increased when the temperature difference issmaller than the predetermined temperature difference.
 12. The system ofclaim 1, wherein the predetermined temperature difference is set basedon a temperature difference between the refrigerant and the hot watergenerating fluid in a vicinity of an intermediate portion of saidradiator.
 13. The system of claim 12, wherein the predeterminedtemperature difference set based on a temperature difference between therefrigerant and the hot water generating fluid in a vicinity of anintermediate portion of said radiator is corrected based on at least oneof the temperature of the hot water generating fluid flowing into saidradiator, the temperature of the hot water generating fluid flowing fromsaid radiator and the refrigerant temperature on a low-pressure side.